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United States Patent |
5,524,659
|
Takata
,   et al.
|
June 11, 1996
|
Flow control valve and control method therefor
Abstract
The present invention provides a flow control valve which makes it possible
to prevent the leakage of hydraulic fluid when the valve is closed while
continuously controlling the flow rate. An axially slidable piston (12)
housed in a housing (11) provided with a first port (13) and second port
(16) is urged by an electromagnet (20). A first closing valve (14) for
connecting and isolating the first port (13) and first fluid chamber (A),
with a valve seat portion (14C) opened opposite one end portion of the
piston (12) in the first fluid chamber (A) and a valve body (14B) which is
seated in and separated from the valve seat portion in accordance with the
axial displacement of the piston (12), is provided between the first port
(13) and the first fluid chamber (A). An orifice (15) is provided in the
conduit connecting the first fluid chamber (A) and the second fluid
chamber (B).
Inventors:
|
Takata; Koji (Itami, JP);
Hashida; Koichi (Itami, JP)
|
Assignee:
|
Sumitomo Electric Industries, Ltd. (Osaka, JP)
|
Appl. No.:
|
430856 |
Filed:
|
April 28, 1995 |
Foreign Application Priority Data
| Jul 13, 1992[JP] | 4-185043 |
| Jun 07, 1993[JP] | 5-135938 |
Current U.S. Class: |
137/8; 137/486 |
Intern'l Class: |
G05D 007/06 |
Field of Search: |
137/8,486,501,504
|
References Cited
U.S. Patent Documents
2272684 | Feb., 1942 | Vickers | 137/501.
|
3502100 | Mar., 1970 | Jonson | 137/501.
|
3729018 | Apr., 1973 | Butterfield | 137/501.
|
4206781 | Jun., 1980 | Salter | 137/504.
|
5018797 | May., 1991 | Takata.
| |
5127435 | Jul., 1992 | Takata | 137/596.
|
5137339 | Aug., 1992 | Brown.
| |
5190068 | Mar., 1993 | Philbin | 137/486.
|
5215357 | Jun., 1993 | Brown.
| |
5261731 | Nov., 1993 | Togo et al.
| |
Foreign Patent Documents |
0369412 | May., 1990 | EP.
| |
0441343 | Aug., 1991 | EP.
| |
3412351 | Oct., 1985 | DE.
| |
4121470 | Jan., 1992 | DE.
| |
487668 | Dec., 1953 | IT | 137/501.
|
972771 | Oct., 1964 | GB | 137/501.
|
2106613 | Feb., 1983 | GB | 137/505.
|
9203319 | Mar., 1992 | WO.
| |
Primary Examiner: Nilson; Robert G.
Attorney, Agent or Firm: Greenblum & Bernstein
Parent Case Text
This application is a division of application Ser. No. 08/196,108, filed
Feb. 23, 1994, now U.S. Pat. No. 5,460,199 which is the National Phase of
PCT/JP93/00948 having an International filing date of Jul. 9, 1993.
Claims
What is claimed is:
1. A control method for a flow control valve comprising:
providing a flow control valve comprising a housing provided with at least
a first port and a second port, a piston housed in the housing so as to be
axially slidable, a first fluid chamber formed between one end portion of
the piston and the housing, a second fluid chamber connected to the second
port formed between the other end portion of the piston and the housing, a
conduit connecting the first fluid chamber and the second fluid chamber,
an orifice provided in the conduit, urging means for axially urging the
piston, and a first closing valve disposed between the first port and the
first fluid chamber, with a valve seat portion opened opposite one end
portion of the piston in the first fluid chamber and a valve body which is
seated in and separated from the valve seat portion in accordance with the
axial displacement of the piston, for connecting and isolating the first
port and the first fluid chamber, wherein hydraulic force acting on the
piston due to differential pressure between the first fluid chamber and
the second fluid chamber generated by the flow rate through the orifice
acts in a direction so as to seat the valve body of the first closing
valve in the valve seat portion;
providing differential pressure detection means for detecting differential
pressure between the first fluid chamber and the second fluid chamber; and
controlling at least one of either the urging force of the urging means
against the piston or the flow-rate sectional area of the orifice in
accordance with a controlled variable found as a function of the
differential pressure detected by the differential pressure detecting
means and the desired flow rate.
2. A control method according to claim 1, further comprising:
providing a third port in the housing; and
axially displacing the piston to open and close the third port to the first
fluid chamber.
3. A control method according to claim 1, wherein the urging means
comprises an electromagnet and said controlling the urging force of the
urging means comprises controlling an amount of electric current to the
electromagnet to vary the urging force thereof.
4. A control method according to claim 1, further comprising:
providing a force applicator for applying a force in opposition to the
urging force to maintain the piston in an initial position; and
setting the opposition force of the force applicator to a minimum force
required to maintain the piston in the initial position when no urging
force is applied by said urging means.
5. A control method according to claim 1, wherein the first closing valve
is a poppet valve having an urging member, said method further comprising:
elastically urging the valve body in a direction such that the valve body
is seated in the valve seat portion.
6. A control method according to claim 1, further comprising:
providing a bypass conduit parallel to the conduit and connecting the first
fluid chamber and the second fluid chamber;
providing a relief valve in the bypass conduit; and
relieving the differential pressure between the first fluid chamber and
second fluid chamber by opening the relief valve when the differential
pressure exceeds a predetermined value.
7. A control method according to claim 6, further comprising:
providing a second bypass conduit parallel to the conduit and connecting
the first fluid chamber and the second fluid chamber;
providing a relief valve in the second bypass conduit; and
relieving the differential pressure between the second fluid chamber and
the first fluid chamber by opening the relief valve when the pressure in
the second fluid chamber exceeds the pressure in the first fluid chamber
by a second predetermined value.
8. A control method for a flow control valve comprising:
providing a flow control valve comprising a housing provided with at least
a first port and a second port, a piston housed in the housing so as to be
axially slidable, a first fluid chamber formed between one end portion of
the piston and the housing, a second fluid chamber connected to the second
port formed between the other end portion of the piston and the housing, a
conduit connecting the first fluid chamber and the second fluid chamber,
an orifice provided in the conduit, urging means for axially urging the
piston, and a first closing valve disposed between the first port and the
first fluid chamber, with a valve seat portion opened opposite one end
portion of the piston in the first fluid chamber and a valve body which is
seated in and separated from the valve seat portion in accordance with the
axial displacement of the piston, for connecting and isolating the first
port and the first fluid chamber, wherein hydraulic force acting on the
piston due to differential pressure between the first fluid chamber and
the second fluid chamber generated by the flow rate through the orifice
acts in a direction so as to seat the valve body of the first closing
valve in the valve seat portion;
providing differential pressure detection means for detecting differential
pressure between the first fluid chamber and the second fluid chamber; and
controlling the urging force of the urging means against the piston in
accordance with a controlled variable found as a function of the
differential pressure detected by the differential pressure detecting
means and the desired flow rate.
9. A control method according to claim 8, wherein the urging means
comprises an electromagnet and said controlling the urging force of the
urging means comprises controlling an amount of electric current to the
electromagnet to vary the urging force thereof.
10. A control method according to claim 8, further comprising:
providing a bypass conduit parallel to the conduit and connecting the first
fluid chamber and the second fluid chamber;
providing a relief valve in the bypass conduit; and
relieving the differential pressure between the first fluid chamber and
second fluid chamber by opening the relief valve when the differential
pressure exceeds a predetermined value.
11. A control method according to claim 10, further comprising:
providing a third port in the housing;
axially displacing the piston to open and close the third port to the first
fluid chamber;
providing a second bypass conduit parallel to the conduit and connecting
the first fluid chamber and the second fluid chamber;
providing a relief valve in the second bypass conduit; and
relieving the differential pressure between the second fluid chamber and
the first fluid chamber by opening the relief valve when the pressure in
the second fluid chamber exceeds the pressure in the first fluid chamber
by a second predetermined value.
12. A control method according to claim 11, further comprising:
providing a groove adjacent the third port in the housing;
providing a seal in the groove; and
sealing the piston with the housing upon the axially displacing the piston
to close the third port to the first fluid chamber.
Description
FIELD OF APPLICATION IN INDUSTRY
The present invention relates to a flow control valve and control method
therefor, and makes possible both the prevention of leakage of hydraulic
fluid when the valve is closed and continuous flow control in a flow
control valve for a car-borne hydraulic apparatus for steering, suspension
or the like, and especially in a flow control valve suitable for use in
the brake hydraulic control system of an anti-lock hydraulic control
system or such like.
DESCRIPTION OF THE RELATED ARTS
Various conventional flow control valves suitable for the hydraulic
apparatus of a brake hydraulic control system have been provided, and the
present applicant, for example, proposes a flow control valve of this kind
in Japanese Laid-Open Patent Application Nos. 3-90462, 3-223578 and
3-234987.
FIG. 8 shows an example of the flow control valve described in Japanese
Laid-Open Patent Application No. 3-90462.
In this flow control valve, a piston 2 is freely slidably disposed in a
housing 1, and a first fluid cheer A and a second fluid chamber B are
respectively provided between each end portion of the piston 2 and housing
1. A conduit 2B with an orifice 2A is provided axially in the piston 2,
and a surface passage 2C connected to the conduit 2B is provided in the
circumferential surface of the piston 2. In addition, the piston 2 is
urged in an axial direction by a spring 3 and electromagnet 4.
Furthermore, an inlet port 1A connected to the conduit 2B by the surface
passage 2C and an outlet port 1B connected to the second liquid chamber B
are provided in the housing 1.
The flow rate through the orifice 2A is determined by the differential
pressure on either side of the orifice connecting the first fluid chamber
A and second fluid chamber B, and the differential pressure is determined
by the urging force of the spring 3 and electromagnet 4 against the piston
2. A closing valve which connects and isolates the inlet port 1A and
surface passage 2C is formed by the sliding of the piston 2 in the housing
1, and the differential pressure on either side of the orifice 2A is
maintained by this closing valve. Consequently, the flow rate of hydraulic
fluid flowing from the inlet port 1A to the outlet port 1B can be
continually adjusted in the flow control valve by adjusting the electric
current applied to the electromagnet 4 so as to change the urging force
that acts against the piston 2.
In addition, although in this flow control valve the fluid pressure of the
outlet port 1B acts in the direction of the axis of the piston 2, the
fluid pressure of the inlet port 1A acts via the closing valve formed by
the inlet port 1A and the surface passage 2C formed on the circumferential
surface of the piston 12, and so does not exert an axial force against the
piston 2. The following Formula (1) for differential pressure Pa-Pb on
either side of the orifice 2A maintained by the action of the closing
valve is thus formed, where Pa is the pressure of the first liquid chamber
A, Pb is the pressure of the second liquid chamber B, f is the urging
force of the spring, F is the urging force of the electromagnet, and A is
the sectional area of the piston:
Pa-Pb=(f-F)/A (1)
As Formula (1) indicates, differential pressure Pa-Pb on either side of the
orifice 2A bears no relation to the fluid pressure of the inlet port 1A
and outlet port 1B in the flow control valve, and is determined solely by
the urging forces f and F, and the sectional area A. A flow rate
proportional to the square root of the differential pressure passes
through the orifice 2A. Namely, the flow control valve has an effect
(pressure compensation effect) which makes it possible to control the flow
rate regardless of input fluid pressure and output fluid pressure.
When the urging force F of the electromagnet 4 is set higher, the discharge
port 1C in FIG. 8 forms a closing valve with the surface passage 2C, and
hydraulic fluid with a flow rate corresponding to the urging force F flows
from the inlet port 1A to the discharge port 1C.
However, as in the above conventional flow control valve a closing valve is
formed by the surface passage 2C in the piston 2 and the inlet port 1A in
the housing 1 so as to obtain a pressure compensation effect, there is a
leakage channel, shown by X in FIG. 8, which connects the inlet port 1A
and outlet port 1B via the sliding surface of the piston 2, even when the
closing valve is in a closed state (i.e. when the inlet port 1A and
surface passage 2C are not connected). Although the channel is extremely
small, it is impossible to prevent the leakage of hydraulic fluid
therethrough.
This leakage of hydraulic fluid makes it extremely difficult to apply this
flow control valve to uses which require that the slightest leakage when
the valve is closed be prevented, such as the discharge valve in an
anti-lock brake system.
An addition to the flow control valve with a pressure compensation effect,
an on/off-type electromagnetic switching valve which only works fully open
and fully closed has been provided. The problem of hydraulic fluid leakage
when the valve is closed does not arise with this electromagnetic
switching valve, but there are limits to the degree to which smooth and
continuous flow control and fine control can be effected, as only full
opening and full closing are effected.
SUMMARY OF THE INVENTION
The object of the present invention is to solve the aforementioned problems
with conventional flow control valves, and to provide a flow control valve
which can continuously control the flow rate without hydraulic fluid
leakage occurring when the valve is closed.
In accomplishing these and other objects of the present invention, there is
provided a flow control valve comprising: a housing provided with at least
a first port and a second port; an axially slidable piston housed in the
housing; a first fluid chamber formed between one end portion of the
piston and the housing; a second fluid chamber formed between the other
end portion of the piston and the housing, and connected to the second
port; a conduit connecting the first fluid chamber and second fluid
chamber; an orifice provided in the conduit; an urging means for urging
the piston in an axial direction; and a first closing valve disposed
between the first port and first fluid chamber, with a valve seat portion
opened in the first fluid chamber opposite one end portion of the piston
and a valve body which is seated in and separated from the valve seat
portion in accordance with the axial displacement of the piston, for
connecting and isolating the first port and first fluid chamber, wherein
the hydraulic force acting on the piston due to the differential pressure
generated between the first fluid chamber and second fluid chamber by the
flow rate passing through the orifice acts in such a direction that the
valve body of the first closing valve is seated in the seating portion.
A poppet valve with a spherical valve body and a seating spring which
elastically urges in such a direction that the valve body is seated in the
valve seat portion is best used for the first closing valve.
When a third port is disposed in the housing and the first closing valve is
in a closed state, a second closing valve which opens and closes in
accordance with the axial displacement of the piston may be provided for
connecting and isolating the third port and first fluid cheer.
A bypass conduit connecting the first fluid chamber and second fluid
chamber may be provided parallel to the conduit, together with a relief
valve in the bypass conduit which opens when differential pressure between
the first fluid chamber and second cheer exceeds a set value.
It is preferable that at least one of the urging force of the urging means
against the piston or the flow-rate sectional area of the orifice be
adjustable.
In addition, a differential pressure detecting means for detecting the
differential pressure between the first port and second port is provided
in the flow control valve according to the present invention. The present
invention provides a control method for a flow control valve according to
which at least one of the urging force of the urging means against the
piston or the flow-rate sectional area of the orifice is controlled in
accordance with the controlled variable found as a function of the
differential pressure detected by the differential pressure detecting
means and the desired flow rate.
By constructing the flow control valve of the present invention as
described above, it is possible to prevent the leakage of hydraulic fluid
when the first closing valve is closed, as the first closing valve which
connects and isolates the first port and first fluid chamber is provided
with a valve body which sits in and separates from the valve seat portion
in accordance with the axial displacement of the piston. It is also
possible to continuously control the rate of flow of hydraulic fluid
between the first port and second port by adjusting the urging force of
the urging means against the piston and the flow-rate sectional area of
the orifice.
In addition, when a third port is provided in a housing with a second
closing valve for connecting and isolating the third port and first fluid
chamber in accordance with the displacement of the piston, it is possible
to not only control the flow rate between the first port and second port,
but also to control the flow rate between the third port and the first
port or second port. The flow control valve according to the present
invention can thus be used as a "dual-effect" flow control valve.
Furthermore, it is possible to increase the flow rate when a bypass conduit
and relief valve are provided, as hydraulic fluid flows through the bypass
conduit without going through the orifice when the differential pressure
between the first fluid chamber and second fluid chamber exceeds a set
value.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a side sectional view showing a flow control valve in a, first
embodiment of the present invention.
FIG. 2 is a side sectional view showing a flow control valve in a second
embodiment of the present invention.
FIG. 3 is a schematic of an anti-lock brake system using the flow control
valve according to the present invention.
FIG. 4 is a schematic of an electronic control brake booster using the flow
control valve according to the present invention.
FIG. 5 is a side sectional view showing a dual-effect type hydraulic
control system in a third embodiment of the present invention.
FIG. 6 is a side sectional view showing a modification of the third
embodiment.
FIG. 7 is a side sectional view showing a flow control valve in a fourth
embodiment of the present invention.
FIG. 8 is a side sectional view showing an example of a conventional flow
control valve.
FIG. 9 is a schematic view of control elements used in controlling the flow
according to the proposed invention.
PREFERRED EMBODIMENTS OF THE PRESENT INVENTION
The following is a detailed description of the present invention based on
the embodiments shown in the drawings, although the present invention is
not thus limited to these embodiments.
First Embodiment
In the flow control valve 10-1 in a first embodiment of the present
invention shown in FIG. 1, an axially slidable piston 12 is housed in an
almost liquidtight state within a housing 11 provided with a first port
13, which forms an input opening, and a second port 16, which forms an
output opening. In addition, a first fluid chamber A is formed between one
end portion of the piston 12 and the housing 11, and a second fluid
chamber B connected to the second port 16 is formed between the other end
portion of the piston 12 and the housing 11.
A conduit 12A connecting the first fluid chamber A and second fluid chamber
B to each other by the orifice 15 is provided through the piston 12. A
pushrod portion 12B for opening the closing valve 14 described below is
disposed protruding from the end portion of the piston 12 on the side of
the first fluid chamber A, and a piston drive element 12C is disposed
protruding from the end portion of the piston 12 on the side of the second
fluid chamber B.
A spring 18 for maintaining the piston 12 in its initial position by urging
the piston 12 axially rightwards in the figure is disposed contracted in
the first fluid chamber A, and an electromagnet 20 which constitutes the
urging means is disposed on the side of the second fluid chamber B in the
housing 11. The piston drive element 12C is inserted into the
electromagnet 20 so that when an electric current is applied, the piston
12 is urged axially leftwards in the figure in accordance with the
electric current. When the electric current applied to the electromagnet
20 is 0 and no urging force is applied against the piston 12 by the
electromagnet 20, the piston 12 is maintained in its initial position by
the urging force of the spring 18, as shown in FIG. 1.
As the urging force of the spring 18 diminishes the output of the
electromagnet 20, the urging force should be set at the minimum required
to maintain the piston 12 in its initial position when the electromagnet
20 is not on.
A first closing valve 14 for connecting and isolating the first port 13 and
first fluid chamber A is provided between the first port 13 and first
fluid chamber A. The first closing valve 14 is a poppet valve comprising a
seating spring 14A, a spherical valve body 14B, and a valve seat portion
14C opened in the first fluid chamber A opposite the end portion of the
piston 12.
The valve body 14B is elastically pressed from the left in the figure
rightwards into the valve seat portion 14C by the seating spring 14A, so
that when in this closed state, the first port 13 and the first fluid
chamber A are isolated from each other. In addition, the valve body 14B is
pushed leftwards in the figure by the pushrod portion 12B when the piston
12 is displaced leftwards in the figure so as to resist the urging force
of the seating spring 14A and open the valve. The first port 13 and the
first fluid chamber A are thus connected.
A bypass conduit 31 connecting the first fluid chamber A and the second
fluid chamber B to each other is disposed in the housing 11 parallel to
the conduit 12A. A relief valve 33 is provided in the bypass conduit 31.
The relief valve 33 comprises a spherical valve body 33B which is
elastically urged leftwards in the figure by the seating spring 33A
against the valve seat portion 33C. When the fluid pressure of the first
fluid chamber A exceeds the fluid pressure of the second fluid chamber B
by a set value, the valve body 33B resists the urging force of the seating
spring 33A and separates from the valve seat portion 33C, thus opening the
valve.
In the flow control valve in the first embodiment, where the effective
sectional area of the seat portion of the first closing valve 14 is a, the
sectional area of the piston 12 is A, the fluid pressure of the first port
13 is Pi, the fluid pressure of the first fluid chamber A is Pa, the fluid
pressure of the second fluid chamber B and second port 16 is Pb, the
urging force of the electromagnet 20 is F (disregarding the very weak
urging force of the seating spring for maintaining the valve body of the
check valve 14), the piston is displaced axially leftwards so that the
closing valve 14 opens when the rightward axial force acting on the piston
12 is greater than the leftward axial force. Accordingly, the closing
valve 14 opens when Formula (2) is formed:
F-f+Pb.times.A-Pa.times.(A-a)>Pi.times.a (2)
A transformation of Formula (2) produces:
(Pa-Pb)<(F-f-(Pi-Pb).times.a)/(A-a) (3)
When Formula (3) is formed, the piston 12 is displaced leftwards in the
figure, and the first closing valve 14 is opened by the pushing motion of
the pushrod 12B. Hydraulic fluid flows from the first port 13 to the first
fluid chamber A via the orifice 15 due to the opening of the first closing
valve. The fluid pressure Pa of the first fluid chamber A is increased by
the flow of hydraulic fluid through the first closing valve 14, and
hydraulic force acting rightwards in the figure (in a direction so as to
seat the valve body 14B) thus acts on piston 12 by means of the
differential pressure between the first fluid chamber A and second fluid
chamber B.
When fluid pressure Pa increases, Formula (3) is no longer formed, and the
piston 12 is displaced axially rightwards so that the first closing valve
14 closes, and the inlet of hydraulic fluid from the first port 13 is cut
off. However, as the first fluid chamber A and second fluid chamber B are
mutually connected by the orifice 15, hydraulic fluid which has flowed
into the first fluid chamber A is discharged to the second fluid chamber B
in proportion to the square root of (Pa-Pb) corresponding with the
flow-rate sectional area of the orifice 15, and the fluid pressure Pa of
the first fluid chamber A falls. Consequently, Formula (3) is once again
formed, and the first closing valve 14 opens so that hydraulic fluid is
again allowed to flow in from the first port 13 to the first fluid chamber
A.
The inlet quantity of hydraulic fluid to the first fluid chamber A and
outlet quantity of hydraulic fluid from the first fluid chamber A are
equalized by this opening/closing action of the first closing valve 14,
and flow is stabilized when the inequality of Formula (3) becomes an
equality. Consquently, if the size of the urging force F of the
electromagnet 20 is adjusted and the right of Formula (3) (F
-f-(Pi-Pb).times.a)/(A-a) set at a certain value, then this value becomes
the pressure differential (Pa-Pb) on the left of Formula (3), and a flow
rate proportional to the square root of the differential pressure (Pa-Pb)
flows from the first port 13 to the second port 16.
When F=0, i.e. When the electric current is off, the right of the Formula
(3) becomes negative, and a state in which Pa=Pb and the flow rate is 0 is
maintained without the force to open the first closing valve 14 being
generated. Namely, the flow control valve 10-1 in the first embodiment is
of a normally closed type wherein the first closing valve 14 is normally
in a closed state when the current to the electromagnet 20 is off.
When the valve is closed, there is no leakage of hydraulic fluid from the
first port 13 and flow to the second port 16, as a popper valve wherein
the valve body 14B is pushed into the seat portion 14C so as to close the
valve is used for the first closing valve 14.
Next, differential pressure (Pa-Pb) increases when F is increased from F=0,
thus increasing the flow rate of hydraulic fluid flowing from the first
port 13 to the second port 16. Accordingly, it is possible to continuously
control the through flow-rate of hydraulic fluid in the first embodiment
by adjusting the electric current applied to the electromagnet 20.
When differential pressure (Pa-Pb) reaches a set value, in accordance with
the increase in the urging force F, the relief valve 33 opens. Hydraulic
fluid which has entered from the first port 13 thus flows to the output
opening 16 via the bypass conduit 31, and the flow rate of the flow
control valve 10-1 temporarily increases.
Of the parameters on the right of Formula (3), values f, A and a are fixed
values determined at the design stage. However, Pb and Pi differ depending
on the apparatuses connected to the first port 13 and second port 16.
Accordingly, the desired flow rate, i.e. the urging force F corresponding
to the pressure differential Pa-Pb, is a function of the differential
pressure between the first port 13 and second port 16.
When, for example, fluid pressure Pi is fixed, the urging force F for
obtaining the desired flow rate is determined by measuring fluid pressure
Pb, and conversely, when fluid pressure Pb is fixed, the urging force F
for obtaining the desired flow is determined by measuring fluid pressure
Pi.
Accordingly, fluid pressures Pb and Pi must be measured and the urging
force F set in accordance therewith in order to obtain the desired flow
rate in the flow control valve 10-1.
Second Embodiment
The following is a description of a second embodiment of the present
invention shown in FIG. 2.
The electromagnet 20 in the flow control valve 10-2 shown in FIG. 2 is of a
retracting type which generates an axial urging force rightwards against
the piston 12 when on.
Further, a spring 18 for maintaining the piston 12 in its initial position
is provided in the second fluid chamber B in the second embodiment, and
the piston 12 is urged leftwards in the direction of the axis by means of
the spring 18.
The conditions under which the closing valve 14 opens in the second
embodiment are therefore provided by the following Formula (4), which
corresponds to Formula (3) in the first embodiment:
(Pa-Pb)<(-F+f-(Pi-Pb).times.a)/A-a) (4)
When the urging force f of the spring 18 in the flow control valve 10-2 is
set so as to exceed Pi.times.a, and the electromagnet 20 is off so that
F=0, Formula (4) is normally formed and the closing valve 14 is in an open
state, so that not only does the fluid pressure Pa of the first fluid
chamber A increase until Pa=Pi, but the fluid pressure Pb of the second
fluid chamber B increases until Pb=Pa=Pi. The flow control valve 10-2 in
the second embodiment is thus of a normally closed type wherein the
closing valve 14 is normally open when the electromagnet 20 is off.
When the electromagnet 20 is on and an urging force F applied to the piston
12, the inlet rate to the first fluid chamber A and the outlet rate are
equalized by the opening and closing of the closing valve 14, as in the
first embodiment, and the flow rate is stabilized when the inequality of
Formula (4) becomes an equality. In addition, the flow rate through the
orifice 15 at this time is proportional to the square root of the
differential pressure (Pa-Pb) of the first fluid chamber A and the second
fluid chamber B. Consequently, if the urging force F of the electromagnet
20' is increased, differential pressure (Pa-Pb) accordingly falls, and the
flow rate of hydraulic fluid flowing from the first port 13 to the second
port 16 is reduced. Thus the flow control valve 10-2 in the second
embodiment, too, can continuously control the through flow-rate by
adjusting the electric current applied to the electromagnet 20.
Furthermore, there is no leakage of hydraulic fluid from the first port 13
to the second port 16 when the urging force F is set larger so that the
closing valve 14 is in a closed state, as the first closing valve 14 is
constructed so that the spherical valve body 14B is pushed into the valve
seat portion 14C.
Further description is omitted as the construction and action of the second
embodiment is in all other respects identical to that of the first
embodiment.
FIG. 3 shows a highly typified example of a commonly known volume expansion
type anti-lock system, wherein the flow control valve 10-2 in the second
embodiment is used as a pressure valve, and the flow control valve 10-1 in
the first embodiment is used as a discharge valve.
In this anti-lock system, a freely slidable deboost piston 42 is
accommodated in a control chamber 41. There is thus a liquidtight
partitioning of the control chamber 41 into a first portion 41A and a
second portion 41B by the deboost piston 42.
An isolation valve 45 for connecting and isolating the master cylinder 43
and wheel brake 44 is disposed in the first portion 41B. The isolation
valve 45 is opened and closed by a pushrod portion 42A disposed on the tip
of the deboost piston 42.
An actuator 48 for storing high-pressure hydraulic fluid emitted from the
pump 47 is connected to the second portion 41B via the pressure valve
consisting of the normally open type flow control valve in the second
embodiment. In addition, a reducing valve consisting of the normally
closed type flow control valve 10-1 in the first embodiment is connected
to the second portion 41B.
The first port 13 in the flow control valve 10-2 forming the pressure valve
is connected to the accumulator 48, and the second port 16 therein is
connected to the second portion 41B. In addition, the first port 13 in the
flow control valve 10-1 forming the reducing valve is connected to the
second portion 41B, and the second port 16 therein is connected to the
reserver 49. Hydraulic fluid from the reserver 49 is pressurized by the
pump 47 and sent to the accumulator 48.
The anti-lock system has, in addition, an electronic control device (not
shown in the figure) for detecting signs of wheel lock. An electric
current is applied to the electromagnets 20 in the respective flow control
valves 10-1, 10-2 in accordance with commands from the electronic control
device.
In order to determine the urging forces of the electromagnets 20
corresponding to the desired flow rates in flow control Valves 10-1 and
10-2 described above, it is necessary to know the respective differential
pressures of the fluid pressures Pi of the first ports 13 and the fluid
pressures Pb of the second fluid chamber B and second port 16. In the case
of the flow control valve 10-2 forming the pressure valve in the system
shown in FIG. 3, however, fluid pressure Pi is a fixed pressure almost
equal to the fluid pressure of the accumulator 48, as the first port 13 is
connected to the accumulator 48, in which pressure is almost fixed. In
addition, fluid pressure Pb in the flow control valve 10-1 forming the
reducing valve is normally almost 0, as the second port 16 therein is
connected to the reserver 49, in which pressure is practically zero.
Accordingly, only fluid pressure Pb for the flow control valve 10-1 and
fluid pressure Pi for the flow control valve 10-2 need be known in order
to determine the urging forces F of the electromagnets 20 for the desired
flow rates in each of the flow control valves 10-1 and 10-2.
Pressure at point C in the system shown in FIG. 3 is therefore measured
directly by a pressure sensor 70, shown schematically in FIG. 9 which
constitutes the differential pressure detecting means. The electronic
control device (i.e., urge force controlling means 71) applies an electric
current to the electromagnets 20 so that an urging force corresponding to
the desired flow rate for each of the flow control valves 10-1 and 10-2 is
generated in accordance with Formulae (3) and (4), as determined by
controlled variable producing means 72 where the measured values are the
fluid pressure Pb of the flow control valve 10-1 and the fluid pressure Pi
of the flow control valve 10-2.
The following is a description of the action of the anti-lock system shown
in FIG. 3.
Normally, the flow control valve 10-2 constituting the pressure valve is
open and the flow control valve 10-1 constituting the reducing valve is
closed, so that the pressure of the accumulator 48 acts on the second
portion 41B of the control chamber 41. The deboost piston 42 is therefore
in an upward position in the figure and the isolation valve 45 is open, so
that hydraulic fluid is supplied from the master cylinder 43 to the wheel
brake 44 in accordance with the amount of pressure applied to the brake
pedal 50.
Under these normal circumstances, the high-pressure hydraulic fluid of the
accumulator 48 and second portion 41b of the control chamber 41 acts on
the flow control valve 10-1, which is in a closed state, but as the first
closing valve 14 of the flow control valve 10-1 consists of a valve body
14B which is pushed into the valve seat portion 14C so as to close the
valve, the fluid pressure of the wheel brake 44 is steadily increased by
pressing the brake pedal 50 without any leakage of hydraulic fluid.
When anti-lock control is effected, the flow control valve 10-2
constituting the pressure valve closes and the reducing valve 10-1 opens,
so that the hydraulic fluid of the second portion 41B of the control cheer
41 is discharged. When the second portion 41B is decompressed, the deboost
piston 42 moves downwards in the figure so that the isolation valve 45
closes. When the deboost piston 42 moves downwards, the volume of the
first portion 41A is increased, and thus the fluid pressure of the wheel
brake 44 is reduced.
At this point, the desired through flow-rate can be continuously controlled
by adjusting the urging force F of the electromagnet 20, as the discharge
valve is the flow control valve 10-1 in the first embodiment. Accordingly,
it is possible to control the fluid pressure of the wheel brake 44 when
anti-lock control is effected in the anti-lock system shown in FIG. 3 more
smoothly than with the example where an on/off type electromagnetic
switching valve is used for the discharge valve.
It should be noted that in the system shown in FIG. 3, the fluid pressure
of the second portion 41B of the control chamber 41 can be found from the
fluid pressure of the first portion 41A and the amount of pressure on the
brake pedal 51. A pressure sensor may thus be disposed at point D in the
figure so as to indirectly measure the fluid pressure at point C
corresponding to the fluid pressure Pb of the flow control valve 10-1 and
the fluid pressure Pi of the flow control valve 10-2. In addition, fluid
pressure at point C may be estimated without the use of a pressure sensor
from the occurrence of the signs of the wheels locking or the past
performance of the flow control valves 10-1, 10-2 and the brake pedal.
FIG. 4 shows a highly typified example in which the normally closed flow
control valve 10-1 in the first embodiment is used as the pressure valve
of a brake booster, and the normally open flow control valve 10-2 in the
second embodiment is used as a discharge valve. This kind of brake booster
is disclosed in Japanese Laid-Open Patent Application No. 2-73557 filed by
the present applicant.
A partition wall portion 51A is provided on the right end portion, as shown
in the figure, of the piston 51 slidably fitted in the master cylinder 43.
The partition wall portion 51A effects a liquidtight partitioning of the
interior of the booster portion 52 disposed continuous with the master
cylinder 43 into a first portion 52A and a second portion 52B. In
addition, a passage 52B which connects the first portion 52A of the
booster portion 52 with the master cylinder 43 when the piston 51 is in
the initial position shown in the figure is provided in the piston 51.
A fluid accumulating portion 54 for storing hydraulic fluid is connected to
the first portion 52A of the booster portion 52. The second port 16 of the
normally closed flow control valve 10-1 forming the pressure valve and the
first port 13 of the normally open flow control valve 10-2 forming the
discharge valve are each connected to the second portion 52B of the
booster portion 52.
The first port 13 of the flow control valve 10-1 is connected to the
accumulator 48. In addition, the second port 16 of the flow control valve
10-2 is connected to the reserver 49, and hydraulic fluid in the reserver
49 is conveyed to the accumulator 48 by means of the pump 47.
With the construction shown in FIG. 4, as in the example of an anti-lock
system shown in FIG. 3, the hydraulic pressure Pi of the first port 13 of
the normally closed flow control valve 10-1 forming the pressure valve is
almost fixed, as the first port 13 is connected to the accumulator 48. In
addition, the fluid pressure Pb of the second port 16 of the normally
closed flow control valve 10-2 forming the reducing valve is almost fixed,
as the second port 13 is connected to the reserver 49. Thus with the
construction shown in FIG. 4, pressure at point C in the figure is
measured by a pressure sensor (not shown in the figure), and the urging
force F of the electromagnets 20 is adjusted so as to obtain the desired
flow rate for control valves 10-1, 10-2 in accordance with Formulae (3)
and (4), where the measured values are the fluid pressure Pb of the flow
control valve 10-1 and the fluid pressure Pi of the flow control valve
10-2.
The flow control valve 10-1 which forms the pressure valve in the
construction shown in FIG. 4 is normally closed, and the flow control
valve 10-2 which forms the reducing valve is normally open, so that the
piston 51 moves leftwards in the figure in accordance with pressure on the
brake pedal 50 without fluid pressure acting on the second portion 52B of
the booster portion 52, and hydraulic fluid is supplied to the wheel brake
44.
Normally, high pressure hydraulic fluid from the accumulator 48 acts on the
flow control valve 10-1 which forms the pressure valve, but there is no
leakage of hydraulic fluid and no risk of displacement of the piston 51 in
spite of the brake pedal 50 not being applied, as the first closing valve
14 of the flow control valve 10-1 is constructed so that the valve body
14B is pushed into the seat portion 14C so as to close the valve.
When boosting, however, the fluid control valve 10-1 which forms the
pressure valve opens and the flow control valve 10-2 which constitutes the
reducing valve closes. High-pressure hydraulic fluid from the accumulator
48 is thus supplied to the second portion 52B of the booster portion 52,
applying pressure, and the piston 51 is displaced to the left regardless
of pressure on the brake pedal 50, thus applying pressure to the foil
cylinder 43.
Pressurization of the foil cylinder 43 at this time can thus be effected
more smoothly than in the example in which the pressure valve is an on/off
type electromagnetic valve, as the flow control valve 10-1 which forms the
pressure valve can continuously control the through flow-rate by adjusting
the urging force F of the electromagnet 20.
When boosting is over, the flow control valve 10-1 which constitutes the
pressure valve closes and the flow control valve 10-2 constituting the
reducing valve opens, so that the hydraulic fluid of the second portion
52B of the booster 52 is discharged to the reserver 49. At this time,
decompression of the second portion 52B can be effected smoothly as the
flow control valve 10-2 can continuously control the flow rate by
adjusting the urging force F of the electromagnet 20, as when boosting.
Third Embodiment
The following is a description of a third embodiment of the present
invention.
The flow control valve 10-3 in the third embodiment shown in FIG. 5 is a
dual-effect type flow control valve which combines the flow control valves
10-1, 10-2 in the first and second embodiments. The flow control valve
10-3 is constructed basically the same as the flow control valve 10-1
shown in FIG. 1. A third port 17 connected to the first liquid chamber A
and disposed in the housing 11 constitutes the discharge opening, and the
piston 12 is used for the second closing valve comprising a spool valve
for opening and closing the third port 17.
Namely, the first closing valve 14 is closed when the piston 12 is in the
initial position shown in FIG. 5, while the third port 17 is open, thus
connecting the third port 17 to the first liquid chamber A. In addition,
when the piston 12 moves the first distance L1 leftwards in the figure,
the third port 17 is closed, and the connection to the first fluid chamber
A cut off, and when the piston 12 moves the second distance L2 leftwards
in the figure, the closing valve 14 opens and the first port 13 is
connected to the first fluid chamber A.
Further, a second bypass passage 32 is provided parallel to a bypass
passage 31 with a relief valve 33 in the third embodiment. The relief
valve 34 disposed in the bypass passage 32 opens when the fluid pressure
Pb of the second fluid chamber B exceeds the fluid pressure Pa of the
first fluid chamber A by a set value.
In the case of the flow control valve 10-3 in the third embodiment, the
first closing valve 14 is closed when F<f. Thus as in the aforementioned
Japanese Laid-Open Patent Application No. 3-90462, when:
(Pb-Pa)<(f-F)/A (5)
the third port 17 is connected to the first fluid chamber A and fluid
pressure Pa falls, and when Formula (5) is no longer formed, the
connection between the two is cut off. By such means, a rate of flow
almost proportional to the square root of the differential pressure
(Pb-Pa) flows through the fixed orifice 15 from the second fluid chamber B
towards the first fluid chamber A, and an equilibrium is established in
which the inequality of Formula (5) becomes an equality. It is thus
possible in the third embodiment to control the flow rate (discharge
flow-rate) of hydraulic fluid flowing from the second port 16 to the third
port 17 when the first closing valve 14 is closed, regardless of the fluid
pressure of the second port 16 and third port 17, and obtain a pressure
compensation effect, by constructing so that the third port 17 and the
first fluid chamber A are connected and isolated by means of the axial
displacement of the piston 12, and changing the urging force F.
When the urging force F is increased, the piston 12 is further displaced
leftwards in the figure from the first distance L1 from the initial
position in the figure in the interval between F>f and
F>f+(Pi-Pb).times.a, so that the first closing valve 14 closes and the
piston 12 closes the third port 17.
With the flow control valve 10-3 in the third embodiment, there is no
leakage of hydraulic fluid from the first port 13 to the second port 16
and third port 17 when the first closing valve 14 is closed, as the first
closing valve 14 comprises a spherical valve body 14B pressed into the
valve seat portion 14C so as to close the valve, as in the first and
second embodiments.
When the urging force F is further increased and F>f+(Pi-Pb).times.a, the
third port 17 is maintained in a closed state, while the first closing
valve 14 opens and closes in accordance with Formula (3) so that hydraulic
fluid with a flow rate corresponding to the urging force F flows from the
first port 13 to the second port 16, as in the first embodiment. It is
thus possible to continuously control the flow rate by adjusting the
urging force F.
The flow control valve 10-3 in the third embodiment may be substituted for
the flow control valves 10-1, 10-2 in the brake booster shown in FIG. 4.
In this case, the first port 1 of the flow control valve 10-3 is connected
to the accumulator, the second port 16 is connected to the second portion
52B of the booster portion 52, and the third port is connected to the
reserver 49.
The flow control valve 10-4 shown in FIG. 6 is a modification of the flow
control valve 10-3 in the third embodiment. In flow control valve 10-4, a
groove 11A for the insertion of the seal material is provided on the side
of the first fluid chamber A of the housing 11, and an elastic ring
material 26 for preventing leakage is accommodated in the groove 11A.
With the flow control valve 10-3 shown in FIG. 6, there is the possibility
of leakage of hydraulic fluid from the sliding surface X of the piston
between the closing of the third port 17 and the opening of the first
closing valve 14, as the second closing valve which opens and closes the
third port 17 is formed from the third port 17 and the circumferential
surface of the piston 12.
However, with the flow control valve 10-4 shown in FIG. 6, the third port
17 is isolated from the first fluid chamber A as the periphery of the
piston 12 adheres closely to the ring material 26 when the piston 12 moves
axially leftwards from the initial position shown in the figure. The
leakage of hydraulic fluid from the third port 17 when both the third port
17 and the first closing valve 14 are closed is thus so restricted that it
may be practically ignored with regard to the flow rate through the
orifice 15.
The elastic ring material is described in Japan Laid-Open Patent
Application Nos. 2-31198 and 2-29828 filed by the present applicant.
Fourth Embodiment
The following is a description of a fourth embodiment of the present
invention shown in FIG. 7.
A first port 13 connected to the first fluid chamber A, a second port 16
connected to the second fluid chamber B, and a third port 17 connected to
the periphery of the piston 55 are provided in the housing 11 housing a
freely sliding piston 55 in the flow control valve 10-5 in the fourth
embodiment of the present invention shown in FIG. 7.
A cylindrical fluid chamber 55A is provided hollowed in the end portion of
the piston 55 on the left in the figure, and a surface passage 55B is
provided on the peripheral surface of the piston 55. The surface passage
55B is connected to the fluid chamber 55A via the passage 55C.
In addition, the valve body 57 of the first closing valve 56 for opening
and closing the first port 13 is disposed so as to be freely slidable in
the fluid chamber 55A. The valve body 57 comprises a main body 57a which
slides axially in the fluid chamber 55A, and a protruding portion 57b
protruding from the main body 57a, and is fitted with a spherical body on
the tip for closing the first port 13. A spring 58 is provided in a
contracted state between the main body 57a and the end portion of the
fluid chamber 55A so as to elastically urge the valve body 57 axially
leftwards in the figure. A catching portion 55A for resisting the urging
force of the spring 58 and catching the valve body 57 is disposed
protruding radially from the portion of the piston 55 in which the fluid
chamber A is formed. A through hole 57d is provided in the main body 57a
of the valve body 57, and the surface passage 55B is normally connected to
the first port 13 via the passage 55C, fluid chamber 55A and through hole
57d.
The surface passage 55B is connected to and isolated from the third port 17
by the axial displacement of the piston 55, and the piston 55 functions as
the second closing valve for connecting and isolating the third port 13
and first fluid chamber A.
A conduit 59 connecting from the first fluid chamber A to the second fluid
chamber B is provided in the housing 11, and a variable orifice 60 is
provided in the conduit 59. The variable orifice 60 comprises a moving
element disposed so as to be freely slidable in the fluid chamber disposed
in the conduit 59, a spring 62 for elastically urging the moving element
61 rightwards in the figure, and an electromagnet 63 for urging leftwards
in the figure against the moving element 61. It is possible to change the
position of the moving element 61 of the variable orifice 60 by adjusting
the electric current to the electromagnet 63, and thus change the
sectional area (flow-rate sectional area) of the part indicated by S in
the figure.
An electromagnet 20 for inserting a piston drive element 55E as in the
first through fourth embodiments is provided in the end portion on the
right side in the figure of the piston 55.
The flow control valve 10-5 in the fourth embodiment differs from the first
through third embodiments in that the first port 13 is used for the
discharge opening, the third port 16 is used for the input opening, and
the second port 17 is used for the output opening.
When the electromagnet 20 is not on and F=0, the valve body 57 of the first
closing valve 56 is pressured into the valve seat portion 56a by the
urging force of the spring 58, and so the first closing valve 56 is in a
closed state. The third port 17 is connected to the first fluid chamber A
via the surface passage 55B, but when:
Pa-Pb>f/A (6)
then the piston 12 moves axially rightwards in the figure, and the
connection between the two is cut off. At this point, hydraulic fluid with
a flow rate corresponding to differential pressure (Pa-Pb) flows from the
first fluid chamber A towards the second fluid chamber B so that the
pressure of the first fluid chamber A falls, and Formula (6) is again
formed. Differential pressure on either side of the variable orifice 60
reaches an equilibrium at a state in which Formula (6) has become an
equality due to the opening and closing of the third port 17 by the piston
12. Consequently, the flow control valve 10-5 in the third embodiment
functions as a pressure compensation valve in which the third port 17 is
on the high-pressure side and the second port 16 is on the high-pressure
side when F=0. In addition, the flow rate at this time can be continuously
controlled by changing the flow-rate sectional area S of the variable
orifice 60.
When the urging force F of the electromagnet 20 is set in proportion to the
urging force f of the spring 18, the first closing valve 14 is closed, and
the third port 17 is closed by the circumferential surface of the piston
55.
If the urging force F of the electromagnet 20 is made larger than the
urging force f of the spring 18, then the piston 12 moves axially
rightwards in the figure, and the third port 13 is in a state isolated
from the first fluid chamber A. Furthermore, the valve body 57 is caught
by the catching portion 55B and moves rightwards in the figure due to the
movement to the right of the piston 55, and the first closing valve 14
opens.
Pb-Pa<F-f (Pb-Pi)/(A-a) (7)
is a condition for the opening of the first closing valve 14, as in the
first through third embodiments. When Formula (7) is formed, hydraulic
fluid flows from the second port 16 to the first port 13 through the
variable orifice 60 so when the fluid pressure Pa of the first fluid
chamber A falls, the first closing valve is isolated. When the fluid
pressure of the first fluid chamber A rises, Formula (7) is again formed,
and the first closing valve 14 opens and closes.
The flow rate of hydraulic fluid flowing from the second port 16 to the
first port 13 stabilizes when the inequality of Formula (7) becomes an
equality due to the opening and closing action of the first closing valve
14. Thus in the flow control valve in the fourth embodiment, too, it is
possible to set the flow rate from the second port 16 to the first port 13
to the desired flow rate by adjusting the urging force F of the
electromagnet 20 if the fluid pressures Pb and Pi are known, as with the
flow control valves in the first through third embodiments.
In addition, it is possible in the present embodiment to adjust the flow
rate of hydraulic fluid by changing the flow-rate sectional area S of the
variable orifice 60, while leaving the urging force F of the electromagnet
20 fixed as it is, due to the provision of a variable orifice 60 as
described above.
The flow control valve 10-5 in the fourth embodiment may be substituted for
flow control valves 10-1, 10-2 in the anti-lock brake system shown in FIG.
3. In this case, the first port 13 of the flow control valve 10-5 is
connected to the reserver 49, the second port 16 is connected to the
accumulator 48, and the third port 17 is connected to the second portion
41B of the control chamber 41.
Other Modifications
It should be noted that the present invention is not limited to the above
embodiments, and that various modification are possible.
Furthermore, decisions such as whether to use a normally closed valve or a
normally open valve, whether to make the first port the input port or the
discharge opening, whether or not to provide a third port, whether or not
to make the urging force against the piston variable, whether or not to
make the flow-rate sectional area of the orifice variable, whether or not
to provide a bypass channel and relief valve, and whether to combine any
of the above, may be freely made in accordance with the use of the device
using the flow control valve, the properties of the hydraulic fluid, the
limits of the flow rate to be continuously controlled, and other factors.
Moreover, although the urging force against the piston and flow-rate
sectional area of the orifice in the above embodiments are not controlled
in steps, control thereof may be effected with multiple step changes, such
as two-step and three-step changes.
EFFECT OF THE INVENTION
As is clear from the aforegoing description, the flow control valve
according to the present invention comprises a piston urged by an urging
means and slidably housed in a housing with at least a first port and a
second port, wherein first and second fluid chambers are formed in each
end portion of the piston, the second chamber is connected to the second
port, and a first closing valve with a valve seat portion opened opposite
the end portion of the piston in the first fluid cheer and a valve body
which is seated in and separates from the valve seat portion in accordance
with the axial displacement of the piston is disposed between the first
port and first fluid chamber. In addition, an orifice is provided in the
conduit connecting the first fluid chamber and the second fluid cheer. The
flow control valve according to the present invention thus makes it
possible to simultaneously prevent the leakage of hydraulic fluid when the
first closing valve is closed, and continuously control the flow rate of
hydraulic fluid by adjusting the urging force of the urging means against
the piston, using a simple, low-cost design. The flow control valve
according to the present invention is consequently of immense practical
value in systems in which it is necessary to measure load fluid pressure
(at least one of the output fluid pressure or the input fluid pressure of
the hydraulic control system, due to the construction of the system) in
order to obtain the desired flow rate, and is of particular value to
systems in which the measurement of load fluid pressure is indispensable
due to system control requirements.
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